Two-speed full-time transfer case with integrated planetary gearset and synchronized range shift

ABSTRACT

A full-time transfer case is equipped with an integrated planetary gearset assembly and a synchronized range shift mechanism to provide high-range and low-range drive connections. The integrated planetary gearset assembly includes a first gearset which acts as a two-speed reduction unit and a second gearset which acts as an interaxle differential. The synchronized range shift mechanism is arranged to concurrently move two components of the first gearset and is synchronized to permit on-the-move range shifts.

CROSS-REFERENCE TO RELATED APPLICATIONS

[0001] This application is a divisional of U.S. patent application Ser.No. 09/981,164 filed on Oct. 17, 2001.

FIELD OF THE INVENTION

[0002] The present invention relates generally to a power transfersystem for controlling the distribution of drive torque between thefront and rear drivelines of a four-wheel drive vehicle. Moreparticularly, the present relates to a full-time transfer case having atwo-speed gear reduction unit and an interaxle differential integratedinto a planetary gear assembly, a synchronized range shift mechanism forestablishing high-range and low-range drive modes, and a biasing clutchfor controlling interaxle slip between the front and rear drivelines.

BACKGROUND OF THE INVENTION

[0003] Due to growing consumer demand for four-wheel drive vehicles, aplethora of different power transfer systems are currently utilized fordirecting power (i.e., drive torque) to all four wheels of the vehicle.For example, in many “part-time” power transfer systems, a transfer caseis installed between the front and rear drivelines and is normallyoperable in a two-wheel drive mode for delivering drive torque to thedriven wheels. However, the four-wheel drive mode is desired, a modeshift mechanism is selectively actuated by the vehicle operator fordirectly coupling the non-driven wheels to the driven wheels forestablishing a part-time or locked four-wheel drive mode. One example ofa part-time transfer case is disclosed in commonly-owned U.S. Pat. No.4,770,280.

[0004] It is also known to use “on-demand” power transfer systems forautomatically directing power to the non-driven wheels, without anyinput or action on the part of the vehicle operator, when traction islost at the driven wheels. Typically, the on-demand feature isincorporated into the transfer case by replacing the mode shiftmechanism with a clutch assembly that is interactively associated withan electronic control system and a sensor arrangement. During normalroad conditions, the clutch assembly is maintained in a non-actuatedcondition such that drive torque is only delivered to the driven wheels.However, when the sensors detect a low traction condition at the drivenwheels, the clutch assembly is automatically actuated to deliver drivetorque to the non-driven wheels. The amount of drive torque transferredthrough the clutch assembly can be varied as a function of specificvehicle dynamics detected by the sensor arrangement. One example of anon-demand power transfer system is disclosed in commonly-owned U.S. Pat.No. 5,323,871.

[0005] As yet a further alternative, some vehicles are equipped with afull-time power transfer system having a transfer case equipped with acenter differential that functions to permit interaxle speeddifferentiation while transferring drive torque to both the front andrear drivelines. To minimize loss of traction due to wheel slip, manyfull-time transfer cases are also equipped with a clutch assembly forlimiting speed differentiation and biasing the torque transferred acrossthe center differential. For example, full-time transfer cases aredisclosed in commonly-owned U.S. Pat. Nos. 5,697,861 and 5,702,321.

[0006] A significant number of the transfer cases discussed above areequipped with a gear reduction unit and a range shift mechanism operablefor permitting the vehicle operator to choose between high-range andlow-range drive modes. In many instances, the vehicle must be stoppedbefore the transfer case can be shifted between its high-range andlow-range drive modes. Unfortunately, the need to stop the vehicle priorto shifting between the high-range and low-range drive modes isinconvenient, particularly upon encountering road conditions or surfaceterrains where continuation of the vehicle's rolling momentum wouldassist in overcoming the conditions encountered. To alleviate thisinconvenience, some two-speed transfer cases are equipped with asynchronized range shift mechanism from permitting “on-the-move”shifting between the high and low ranges.

[0007] In an effort to minimize the overall size of full-time two-speedtransfer cases, it has been proposed to incorporate the gear reductionunit and the interaxle differential into a common planetary gearassembly. For example, commonly-owned U.S. Pat. No. 5,902,205 disclosesa full-time two-speed transfer case equipped with an integratedplanetary gearset which is operable for establishing full-timehigh-range and low-range four-wheel drive modes through on-the-moveshifting of a synchronized range shift mechanism. While such anarrangement provides a compact construction, there is a continuing needto develop alternatives which meet modern requirements for low noise andweight while advancing the state of the four-wheel drive art.

SUMMARY OF THE INVENTION

[0008] It is therefore an object of the present invention to provide atransfer case for a full-time four-wheel drive vehicle having aplanetary gear assembly which integrates a gear reduction unit and aninteraxle differential into a common arrangement.

[0009] As an additional object of the present invention, the full-timetwo-speed transfer case includes a synchronized range shift mechanismwhich can be selectively actuated for establishing a full-timefour-wheel high-range drive mode, a neutral mode, and a full-timefour-wheel low-range drive mode.

[0010] According to another object of the present invention, thetransfer case includes a biasing clutch which is operably associatedwith the outputs of the planetary gear assembly for limiting speeddifferentiation and regulating the drive torque distributiontherebetween in response to the occurrence of slip between the front andrear output shafts of the transfer case. To this end, a control systemis provided which includes sensors for detecting and generating sensorsignals indicative of various dynamic and operational characteristics ofthe vehicle, and a controller for controlling actuation of the biasingclutch in response to the sensor signals. Upon the occurrence oftraction loss, the clutch is automatically actuated for limitinginteraxle slip while transferring increased drive torque to thenon-slipping driveline.

[0011] According a preferred embodiment of the present invention, theplanetary gear assembly is operably installed between and input shaftand front and rear output shafts of the transfer case and is constructedin a compact arrangement. The planetary gear assembly includes a firstplanetary gearset and a second planetary gearset which areinterconnected by a common carrier assembly. The first planetary gearsetis operably installed between the input shaft and the second planetarygearset for driving the carrier assembly at either of a direct speedratio (i.e., high-range) or a reduced speed ratio (i.e., low-range)relative to the input shaft. The common carrier assembly acts as theinput to the second planetary gearset which has first and second outputsrespectively connected to the rear and front output shafts of thetransfer case. Thus, the second planetary gearset functions as aninteraxle differential for permitting speed differentiation anddistributing drive torque between the front and rear output shafts ofthe transfer case.

[0012] Additional objects come with features and advantages of thepresent invention will become apparent from studying the followingdetailed description and appended claims when taken in conjunction withaccompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

[0013]FIG. 1 is a schematic representation of a four-wheel drive motorvehicle equipped with a full-time power transfer system according to thepresent invention;

[0014]FIG. 2 is a sectional view of a full-time two-speed transfer caseconstructed according to a preferred embodiment of the presentinvention;

[0015]FIG. 3 is an enlarged partial view taken from FIG. 2 showingcomponents of the synchronized range shift mechanism in greater detail;

[0016]FIG. 4 is an enlarged partial view taken from FIG. 2 showing thecomponents of the integrated planetary gear assembly in greater detail;

[0017]FIG. 5 is a sectional view of a full-time two-speed transfer caseconstructed according to an alternative preferred embodiment of thepresent invention;

[0018]FIG. 6 is an enlarged partial view of FIG. 5 showing thesynchronized range shift mechanism and integrated planetary gearassembly in greater detail; and

[0019]FIG. 7 is a partial sectional view showing another alternativepreferred embodiment of a synchronized range shift mechanism andintegrated planetary gear assembly for use in a full-time two-speedtransfer case of the present invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0020] Referring now to the drawings, a drivetrain for a four-wheeldrive vehicle is schematically shown interactively associated with apower transfer system 10 of the present invention. The motor vehicledrivetrain includes a front driveline 12 and a rear driveline 14 bothdriveable from a source of power, such as an engine 16, through atransmission 18 which may be of the manual or automatic type. In theparticular embodiment shown, the drivetrain includes a transfer case 20for transmitting drive torque from engine 16 and transmission 18 tofront driveline 12 and rear driveline 14. Front driveline 12 includes apair of front wheels 22 connected at opposite ends of a front axleassembly 24 having a front differential 26 coupled to one end of a frontdrive shaft 28, the opposite end of which is coupled to a front outputshaft 30 of transfer case 20. Similarly, rear driveline 14 includes apair of rear wheels 32 connected at opposite ends of a rear axleassembly 34 having a rear differential 36 coupled to one end of a reardrive shaft 38, the opposite end of which is interconnected to a rearoutput shaft 40 of transfer case 20.

[0021] As will be detailed hereinafter with greater specificity,transfer case 20 is equipped with a planetary gear assembly 42, a rangeclutch assembly 44, and a mode clutch assembly 46. Planetary gearassembly 42 includes a first planetary gearset 48 and a second planetarygearset 50 interconnected through a common carrier assembly 52. Rangeclutch assembly 44 is operable to couple components of first planetarygearset 48 for driving carrier assembly 52 at either of a first(high-range) speed ratios or a second (low-range) speed ratio. Secondplanetary gearset 50 of planetary gear assembly 42 functions as aninteraxle differential having carrier assembly 52 as its input withother components thereof operably coupled to front output shaft 30 andrear output shaft 40 of transfer case 20. Additionally, mode clutchassembly 48 is operable to control the magnitude of speeddifferentiation and torque biasing between rear output shaft 40 and rearoutput shaft 30. Power transfer system 10 further includes apower-operated actuator assembly 54 for controlling actuation of rangeclutch assembly 44 and mode clutch assembly 46, a sensor group 56 forsensing dynamic and operational characteristics of the motor vehicle andgenerating sensor input signals indicative thereof, and a controller 58for generating control signals in response to the sensor input signals.Power transfer system 10 also includes a mode select mechanism 60 forpermitting the vehicle operator to select one of the available drivemodes. In particular, controller 58 functions to control actuation ofpower-operated actuator assembly 54 in response to the mode signal sentto controller 58 from mode select mechanism 60 that is indicative to theparticular mode selected.

[0022] With particular reference now to FIGS. 2 through 4 of thedrawings, transfer case 20 is shown to include an input shaft 62rotatably supported by a bearing assembly 64 from a housing 66. Inputshaft 62 is adapted for connection to an output shaft of transmission18. Likewise, front output shaft 30 and rear output shaft 40 are alsorotatably supported by housing 66. As noted, planetary gear assembly 42includes a first gearset 48 interconnected to second gearset 50 via acommon carrier assembly 52. First gearset 48 includes a ring gear 68, afirst sun gear 70, and a set of first planet gears 72 each meshed withring gear 68 and first sun gear 70. Planet gears 72 are rotatablysupported on long pins 74 and short pins 76, each of which extendsbetween first and second carrier rings 78 and 80, respectively, ofcarrier assembly 52. First sun gear 70 is shown fixed via a splinedconnection 84 for rotation with and axial sliding movement on a quillshaft 86 journalled on rear output shaft 40 and which is fixed via asplined connection 88 for rotation with input shaft 62. As such, firstsun gear 70 is always driven by input shaft 62.

[0023] Ring gear 68 is shown to include a radial plate segment 90 thatis fixed via bolts 92 to a coupling ring 94. Coupling ring 94 includes acircumferential groove within which a radial lug segment 96 of first sungear 70 is retained. Coupling ring 94 permits first sun gear 70 torotate freely relative to ring gear 68 while also enabling concurrentaxial movement of ring gear 68 and sun gear 70 relative to carrierassembly 52 between a first position (denoted by position line “A”) anda second position (denoted by position line “B”). When ring gear 68 andfirst sun gear 70 are located in the A position, external clutch teeth98 on first sun gear 70 are engaged with internal clutch teeth 100 onsecond carrier ring 80. As such, first sun gear 70 couples carrierassembly 52 for common rotation with input shaft 62. In contrast, sungear clutch teeth 98 are released from engagement with clutch teeth 100on second carrier ring 80 when ring gear 68 and first sun gear 70 arelocated in the B position, thereby allowing carrier assembly 52 torotate relative to input shaft 62.

[0024] Range clutch assembly 44 is shown to include a range shiftmechanism 102 having a first clutch plate 104 fixed (i.e., splined) toinput shaft 62, a second clutch plate 106 fixed to housing 66, a clutchhub 108 journalled on portions of input shaft 62 and quill shaft 86, anda dog-type range sleeve 110 that is splined to clutch hub 108 forrotation therewith and axial movement thereon between three distinctrange positions. A non-synchronized version of range shift mechanism 102is shown in the upper-half of FIG. 2. Preferably, however, range shiftmechanism 102 is synchronized to permit “on-the-move” shifting. Thus, asynchronized version of range shift mechanism 102 is shown in thelower-half of FIG. 2. The synchronized version of range shift mechanism102 includes a first synchronizer 112 disposed between clutch hub 108and first clutch plate 104 and a second synchronizer 114 disposedbetween clutch hub 108 and second clutch plate 106. Clutch hub 108 isshown to include an axially extending shaft segment 116 that isjournalled on quill shaft 86. Ring gear 68 is coupled to clutch hub 108via a splined connection 118 provided between shaft segment 116 ofclutch hub 108 and an axially-extending hub 120 formed on plate segment90 of ring gear 68. Splined connection 118 permits axial movement ofring gear 68 between it's A and B positions while maintaining amechanical coupling with clutch hub 108.

[0025] As noted, range sleeve 110 is moveable between three distinctrange positions shown to include a high-range (“H”) position, alow-range (“L”) position, and a neutral (“N”) position. To this end, afirst range fork 124 is provided which moves under the control ofactuator assembly 54 to control axial movement of range sleeve 110between its three range positions. First range fork 124 has a forksegment retained in a groove formed in range sleeve 110. Likewise, asecond range fork 128 is provided which also moves under the control ofactuator assembly 54 to control axial movement of ring gear 68 and firstsun gear 70 between the two distinct positions. As seen, second rangefork 128 has a fork segment retained in a groove formed in the outerperiphery of ring gear 68. As will be detailed, actuator assembly 54includes a drive mechanism which functions to coordinate movement ofrange forks 124 and 128 so as to coordinate axial movement of rangesleeve 110 and ring gear 68 to establish different drive connectionsbetween input shaft 62 and carrier assembly 52.

[0026] A first or high-range drive connection is established betweeninput shaft 62 and carrier assembly 52 when range sleeve 110 is in its Hposition and ring gear 64 is in its A position, as shown in the upperhalf of FIG. 2. With range sleeve 110 in its H position, its internalclutch teeth and engaged with external clutch teeth on first clutchplate 104 such that clutch hub 108 and ring gear 68 are coupled forcommon rotation with input shaft 62. Concurrently, positioning of ringgear 68 and first sun gear 70 in the A position causes external clutchteeth 98 on first sun gear 70 to engage internal clutch teeth 100 onsecond carrier ring 80. Thus, ring gear 68, first sun gear 70 andcarrier assembly 52 are all commonly driven by input shaft 62. Thisestablishes a direct speed ratio drive connection between input shaft 62and carrier assembly 52 such that first planet gears 72 are unloadedduring operation of transfer case 20 in the high-range modes. This is asignificant advantage over conventional two-speed planetary units sinceit eliminates concerns about fretting corrosion of the teeth on firstplanet gears 72 and permits use of quieter and smaller helical gearsinstead of conventional spur gears.

[0027] A second or low-range drive connection is established betweeninput shaft 62 and carrier assembly 52 when range sleeve 110 is in its Lposition and ring gear 68 is in its B position, as is shown in thelower-half of FIG. 2. With range sleeve 110 in its L position, itsinternal clutch teeth are engaged with external clutch teeth formed onsecond clutch plate 106 such that clutch hub 108 and ring gear 68 arebraked against housing 66 to prevent rotation. In addition, movement ofring gear 68 and first sun gear 70 to the B position causes first sungear 70 to slide axially to a position whereat its clutch teeth 98 aredisengaged from clutch teeth 100 on second carrier ring 80. As such,driven rotation of input shaft 62 drives first sun gear 70 via quillshaft 86 such that carrier assembly 52 is rotatively driven at a reducedspeed due to ring gear 68 being braked. Finally, a neutral (non-driven)mode is established when range sleeve 110 is in its N position and ringgear 68 is in its A position. With range sleeve 110 in its N position,its internal clutch teeth are disengaged from the external clutch teethon first and second clutch plates 104 and 106. In this mode, no drivetorque is delivered from input shaft 62 to carrier assembly 52.

[0028] With continued reference to FIGS. 2 through 4, second gearset 50is shown as a dual-planetary arrangement which functions as an interaxledifferential to permit speed differentiation and distribute drive torquebetween front output shaft 30 and rear output shaft 40. Carrier assembly52, when driven at either of the high-range and low-range speed ratios,acts as the input to second gearset 50 which has its outputs coupled tooutput shafts 30 and 40. In particular, second gearset 50 includes asecond sun gear 136, a third sun gear 138, a set of full-length secondplanet gears 140, and a set of half-length third planet gears 142.Second sun gear 136 is shown to be fixed via a splined connection 144 torear output shaft 40 while third sun gear 138 is journally supportedthereon. Second planet gears 140 are rotatably supported on long pins 74while third planet gears 142 are rotatably supported on short pins 146.Long pins 74 are shown to extend between second carrier ring 80 and athird carrier ring 148 while short pins 146 are shown to extend betweenthird carrier ring 148 and a fourth carrier ring 149. Second planetgears 140 are meshed with second sun gear 136 while third planet gears142 are meshed with third sun gear 138. In addition, second and thirdplanet gears 140 and 142 are circumferentially arranged in meshed pairs.According to the particular construction shown, second sun gear 136 actsas a first output of second gearset 50 while third sun gear 138 acts asthe second output thereof. Third sun gear 138 is fixed to a drivesprocket 150 that is operable for transferring drive torque to a drivensprocket 152 fixed to front output shaft 30. A power chain 154 is shownto interconnect driven sprocket 152 to drive sprocket 150.

[0029] As best seen from FIG. 4, mode clutch assembly 46 is arranged tocontrol speed differentiation and torque biasing between front outputshaft 30 and rear output shaft 40. Mode clutch assembly 46 is amulti-plate friction clutch which includes a clutch hub 160 that isfixed to rear output shaft 40, a clutch drum 162 integrally formed withdrive sprocket 150, and a clutch pack 164 disposed therebetween. Clutchpack 164 includes a set of inner friction plates splined to clutch hub160 and which are interleaved with a set of outer friction platessplined to drum 162. Clutch pack 164 is located between a reaction plate166 formed integrally with clutch hub 160 and an apply plate 168 that issplined to drum 162. As will be detailed, movement of apply plate 168relative to clutch pack 164 functions to vary the compressive engagementforce exerted thereon for adaptively regulating speed differentiationand torque biasing between front output shaft 30 and rear output shaft40.

[0030] Mode clutch assembly 46 further includes a clutch actuationmechanism 170 for moving apply plate 168 under the control of actuationassembly 54. Mechanism 170 includes a locator plate 172 that is splinedfor rotation with clutch drum 162, a pressure plate 174, and a set ofthrust pins 176 having one end fixed to pressure plate 174 and whichextend through apertures in locator plate 172. The second end of thrustpins 176 are adapted to engage apply plate 168. A return spring 178urges pressure plate 174 in a direction away from locator plate 172 fornormally retracting thrust pins 176 from engagement with apply plate168. However, axial movement of pressure plate 174 between afully-retracted position and a fully-extended position causes thrustpins 176 to exert a clutch engagement force on apply plate 168 thatvaries between predetermined minimum and maximum values.

[0031] To provide means for moving pressure plate 174 between its fullyretracted and fully extended positions, clutch actuation mechanism 170includes a thrust assembly 180, a lever arm 182, and a mode fork 184.Mode fork 184 has a tubular segment 186 fixed to a shift rail 188, theopposite ends of which are supported in sockets formed in housing 66.Thrust assembly 180 includes a thrust ring 194 and a thrust bearingassembly 196 that is disposed between thrust ring 194 and pressure plate174. In addition, lever arm 182 is mounted to a pivot post 198 forpivotal movement relative to thrust assembly 180. Lever arm 182 includesa first end portion 200 that is journalled on shift rail 188 and whichengages one end of mode fork 184. The opposite end portion 202 of leverarm 182 is a C-shaped biforcated section partially surrounding rearoutput shaft 40 and which engages thrust ring 194. In operation, axialmovement of mode fork 184 causes corresponding pivotal movement of leverarm 182 which, in turn, controls movement of thrust assembly 180 andpressure plate 174.

[0032] Preferably, actuator assembly 54 includes a rotary actuator, suchas an electric gearmotor 206, which is operable for generating an outputtorque, the value of which varies as a function of the magnitude of theelectrical control signal applied thereto by controller 60. To providemeans for selectively controlling the magnitude of the clutch engagementforce exerted on clutch pack 164 and coordinate movement of range forks124 and 128, actuator assembly 54 further includes a drive mechanism208. Drive mechanism 208 is interconnected to a rotary output member 210of gearmotor 206 for changing its output torque into an axially-directedforces that are used for controlling axial movement of range forks 124,128 and mode fork 184. According to a preferred construction, drivemechanism 208 includes a sector plate 212 that is rotatably driventhrough a range of angular motion by output member 210 of gearmotor 206.

[0033] To generate axial movement of mode fork 184, sector plate 212includes a mode slot 214 within which a mode pin 216 is retained. Modepin 216 is fixed to a flange section 218 of mode fork 184. The contourof mode slot 214 is configured to cause the desired direction and amountof axial movement of mode fork 184 in response to rotation of sectorplate 212 for generating the desired clutch engagement force exerted byactuation mechanism 170 on clutch pack 164. To control axial movement ofrange sleeve 110, sector plate 212 also has a first range slot 220within which a first range pin 222 extends. First range pin 222 is fixedto a tubular segment 224 of first range fork 124 which is shownsupported for sliding movement on shift rail 188. The contour of firstrange slot 220 is configured to cause controlled axial movement of rangesleeve 110 in response to controlled rotation of sector plate 212. In asimilar fashion, sector plate 212 includes a second range slot 226within which a second range pin 228 extends. Second range pin 228 isfixed to a tubular segment 230 of second range fork 128 which islikewise supported for sliding movement on shift rail 188. Again, thecontour of second range slot 226 is configured to cause control axialmovement of ring gear 68 and first sun gear 70 in response to controlledrotation of sector plate 212.

[0034] According to a preferred embodiment of the present invention,sector plate 212 may be rotated to any one of five distinct sectorpositions to establish a corresponding number of drive modes. Thesemodes may include a locked four-wheel high-range drive mode, a full-timefour-wheel high-range drive mode, a neutral mode, a full-time four-wheellow-range drive mode, and a locked four-wheel low-range drive mode. Theparticular four-wheel drive mode selected is established by the positionof mode pin 216 in mode slot 214, the position of first range pin 222 infirst range slot 220, and the position of second range pin 228 andsecond range slot 226. In operation, the vehicle operator selects adesired four-wheel drive mode via actuation of mode select mechanism 60which, in turn, sends a mode signal to controller 58 that is indicativeof the selection. Thereafter, controller 58 generates an electriccontrol signal that is applied to gearmotor 206 for controlling therotated position of sector plate 212. More particularly, upon selectionof the locked four-wheel high-range drive mode, the neutral mode, or thelocked four-wheel low-range drive mode, sector plate 212 is controllablyrotated to a predefined sector position associated with each mode.However, when either of the full-time four-wheel high-range or low-rangedrive modes are selected, power transfer system 10 is operable formodulating the clutch engagement force applied to clutch pack 164 ofmode clutch assembly 46 as a function of the various sensor inputsignals.

[0035] Mode select mechanism 60 can take the form of any mode selectordevice which is under the control of the vehicle operator for generatinga mode signal indicative of the specific mode selected. In one form, themode selector device may be in an array of dash-mounted push buttonswitches. Alternatively, the mode selector may be a manually-operableshift lever sequentially moveable between a plurality of positionscorresponding to the available operational modes which, in conjunctionwith a suitable electrical switch arrangement, generates a mode signalindicating the selected mode. In either form, mode select mechanism 60offers the vehicle operator the option of deliberately choosing betweenthe various operative drive modes.

[0036] When the locked full-time four-wheel high-range drive mode isselected, sector plate 212 is rotated to a sector position causing rangesleeve 110 to move to its H position, ring gear 68 to move to it's Aposition, and mode fork 184 to move to a position whereat pressure plate174 is in its fully extended position. As such, the maximum clutchengagement force is exerted on clutch pack 164 and mode clutch assembly46 is considered to be operating in a fully actuated (locked-up)condition. Thus, speed differentiation between rear output shaft 40 andfront output shaft 30 is prevented. Power transfer system 10 may alsoinclude a brake which is an electrically controlled device. The brake isengaged once sector plate 212 is rotated to its sector positioncorresponding to the locked full-time four-wheel high-mode for lockingsector plate 212 against further rotation.

[0037] If mode select mechanism 60 thereafter signals selection of thefull-time four-wheel high-range drive mode, gearmotor 206 is actuatedfor initially rotating sector plate 212 to a position causing mode fork184 to move to a position whereat pressure plate 174 is in its fullyretracted position while range sleeve 110 is maintained in its Hposition and ring gear 68 is maintained it its A position. As such, theminimum clutch engagement force is exerted on clutch pack 164 such thatmode clutch assembly 46 is considered to be in a mon-actuated condition,thereby permitting unrestricted speed differentiation between the outputshafts. However, in the full-time four-wheel high-range drive mode, modeclutch assembly 46 provides adaptive control of speed differentiationand torque biasing. Specifically, the actuated state of gearmotor 206 iscontinuously monitored and modulated in accordance with specificpredefined relationships based on the current value of the sensor inputsignals. As is apparent, the magnitude of the clutch engagement force isvaried by bi-directional rotation of sector plate 212 between itsfull-time and locked high-range sector positions.

[0038] Power transfer system 10 also permits transfer case 20 to beshifted into the neutral mode upon mode selection mechanism 60 signalingselection thereof. Controller 58 commands gearmotor 206 to rotate sectorplate to a neutral sector position. In this sector position, the contourof first range slot 220 has caused range sleeve 110 to move to its Nposition while the contour of second range slot 226 has caused ring gear68 to move to, or remain in, it's A position. Likewise, mode slot 214has caused mode fork 184 to move to a position whereat mode clutchassembly 46 is non-actuated.

[0039] If a full-time four-wheel low-range drive mode is made available,its selection would cause gearmotor 206 to rotate sector plate 212 to acorresponding sector position whereat range sleeve 110 is in its Lposition, ring gear 68 is in its B position, and mode clutch assembly isnon-actuated. Again, the contour of the range and mode slots control thecoordinated movement of range forks 124, 128 and mode fork 184 toestablish the desired mode. Preferably, automatic clutch control in thefull-time low-range drive mode is similar to that described thefull-time four-wheel high-range drive mode. To accomplish this adaptiveclutch control, sector plate 212 must be moveable from its full-timelow-range sector position to a locked four-wheel low-range drive modesector position where a maximum engagement force is applied to clutchpack 164. As before, such rotation of sector plate 212 occurs whilerange sleeve 110 is maintained in its L position and ring gear 68 ismaintained in its B position. Automatic control of mode clutch assembly46 is then accomplished in the full-time four-wheel low-range drive modeto bias torque and limit slip automatically. Finally, selection of thelocked four-wheel low-range drive mode signals controller 58 to rotatesector plate 212 to its corresponding sector position. In this sectorposition, range sleeve 110 is in its L position, ring gear 68 is in itsB position, and mode fork 184 is in the position where pressure plate174 is in its fully extended position such that mode clutch assembly 46is locked-up. As before, the brake can be applied to hold sector plate212 in this position so as to allow gearmotor 206 to be turned-off,thereby decreasing its on-time service.

[0040] Referring now to FIGS. 5 and 6, a modified version of transfercase 20 is designated by reference numeral 20A and the same numbers areused to identify common components. Basically, transfer case 20A issubstantially similar in structure and function to that of transfer case20 except that first planetary gearset 48 was arranged to provide aratio of about 2.6 to 1 for its low-range while first planetary gearset48A of transfer case 20A is adapted to provide a ratio of about 3.9 to 1for its low-range. Specifically, in first planetary gearset 48, firstsun gear 70 has 55 teeth while ring gear 68 has 89 teeth and firstplanet gears 72 each have 17 teeth to define the 2.6:1 ratio. Incontrast, first planetary gearset 48A has a first sun gear 70A with 31teeth while ring gear 68A has 89 teeth and first planet gears 72A eachhave 29 teeth. To accommodate the different size requirements for firstsun gear 70A and first planet gears 72A, transfer case 20A includes amodified coupling ring 94A which is still functional to couple sun gear70A for axial movement with ring gear 68A while permitting relativerotation therebetween.

[0041] First sun gear 70A is shown to be fixed via a splined connection84A for axial movement relative to quill shaft 86A. Coupling ring 94A isfixed to plate segment 90A of ring gear 68A and includes a tubularsegment 236 and a radial ring segment 238. Tubular segment 236 ofcoupling ring 94A is fixed via a splined connection 240 to clutch hub108 such that ring gear 68A is coupled for rotation with clutch hub 108while axially moveable relative thereto. Ring segment 238 of couplingring 94A is shown retained in a circumferential groove 242 formed infirst sun gear 70A. Thus, sun gear 70A is again coupled for axialmovement with ring gear 68A between the A and B positions while stillcapable of rotation relative to ring gear 68A. As with transfer case 20,actuator assembly 54 is again provided for controlling coordinatedmovement of range sleeve 110 between its three distinct positions andmovement of ring gear 68A between its two distinct positions toestablish the high-range and low-range drive connections between inputshaft 62 and carrier assembly 52.

[0042] Referring now to FIG. 7, a transfer case 20B is schematicallyshown to be a modified version of transfer case 20 such that commonreference numerals are again used to identify similar components.Basically, transfer case 20B has a modified synchronized range shiftmechanism 260 that combines movement of range sleeve 110 and ring gear68 of transfer case 20 so as to provide reduced complexity and minimizedpackaging requirements. Specifically, planetary gear assembly 42includes a first gearset 48B interconnected to second gearset 50 viacarrier assembly 52. First gearset 48B includes a first sun gear 70B, aring gear 68B and first planet gears 72 rotatably supported betweencarrier rings 78B and 80. Sun gear 70B is shown to have a coupling ring262 fixed thereto which includes an inner cylindrical rim segment 264and an outer cylindrical rim segment 266 interconnected by a platesegment 268. Inner rim segment 264 has a set of first internal splineteeth 270 which are axially offset from from a set of second internalspline teeth 272. Also, input shaft 62B is shown to include a first setof external spline teeth 274 and a second set of external spline teeth276. A radial lug 278 extending outwardly from outer ring segment 266 isretained in a circumferential groove 280 formed in range sleeve 110B.Ring gear 68B is shown to be fixed to one end of range sleeve 110B suchthat it and sun gear 70B are axially moveable with range sleeve 110B.

[0043] Range sleeve 110B is axially moveable between three distinctrange positions (L, N, H) via movement of range fork 124 upon controlledactuation of actuator assembly 54. Range sleeve 110B includes internalclutch teeth 282 which are in constant mesh with external teeth 284formed on a low hub 286. Low hub 286 is rotatably supported on a lowclutch plate 288 that is fixed to housing 66. A low synchronizerassembly 290 is dosposed between low hub 286 and low clutch plate 288and functions to establish speed synchronization therebetween prior topermitting clutch teeth 282 of range sleeve 110B to enter intoengagement with clutch teeth 292 on low clutch palte 288 during movementof range sleeve 110B into its L position. When a four-wheel low-rangedrive mode is selected, gearmotor 206 rotates sector plate 212 of drivemechanism 208 for causing range fork 124 to move range sleeve 110B toits L position Such movement of range sleeve 110B causes both sets ofclutch teeth 270 and 272 on coupling ring 262 to meshingly engagecorresponding sets of clutch teeth 274 and 276 on input shaft 62B whilealso causes its clutch teeth 282 to engage clutch teeth 292 on lowclutch plate 288. Thus, sun gear 70B is driven by input shaft 62B andring gear 68B is braked by housing 66 against rotation such that carrierassembly 52 is driven at a reduced speed ratio. First planetary gearset48B can be arranged to provide any suitable reduction ratio including,without limitation, either of the 2.6:1 or 3.9:1 ratio previouslydisclosed.

[0044] With continued reference to FIG. 7, synchronized range shiftmechanism 260 is further shown to include a high clutch hub 300 that isrotatably supported on coupling ring 262 and which has external splineteeth 302 in constant mesh with internal clutch teeth 304 formed inouter ring segment 266 of coupling ring 262. First carrier ring 78B isshown to include clutch teeth 306 that are aligned to receive clutchteeth 304 of coupling ring 262 upon movement of range sleeve 110B to itsH position. A high synchronizer assembly 308 is disposed between hub 300and carrier ring 78B and functions to establish speed synchronizationbetween carrier assembly 52 and sun gear 70B prior to engagement ofcoupling ring teeth 304 with carrier ring teeth 306. When it is desiredto establish a four-wheel high-range drive mode, range sleeve 110B ismoved to its H position where teeth 270 on coupling ring 262 engageteeth 276 on input shaft 62B such that sun gear 70B is driven by inputshaft 62B. Also, upon synchronization, clutch teeth 304 on coupling ring262 engages clutch teeth 306 on first carrier ring 78B such that adirect drive connection between input shaft 62B and carrier assembly 52is established. Range sleeve 110B is shown in its N position withcoupling ring 262 disengaged from input shaft 62B.

[0045] The foregoing discussion discloses and describes the preferredembodiments for the present invention. However, one skilled in the artwill readily recognize from such discussion, and from the accompanyingdrawings and claims, various changes, modifications and variations canbe made therein without departing from the true spirit and fair scope ofthe invention as defined in the following claims.

What is claimed is:
 1. A transfer case comprising: an input shaft; firstand second output shafts; a planetary gear assembly interconnecting saidinput shaft to said first and second output shafts and including firstand second gearsets having a common carrier assembly, said first gearsetincluding a first sun gear driven by said input shaft, a ring gear, anda first planet gear supported by said carrier assembly and meshed withsaid first sun gear and said ring gear, said second gearset including asecond sun gear connected to said first output shaft, a third sun gearoperably connected to said second output shaft, a second planet gearsupported by said carrier assembly and meshed with said second sun gear,and a third planet gear supported by said carrier assembly and meshedwith said third sun gear and said second planet gear; a range sleevefixed to said ring gear for movement between a high-range position and alow-range position, said first sun gear is retained for movement withsaid range sleeve while permitting relative rotation between said ringgear and said first sun gear, said range sleeve is operable in itshigh-range position to couple said first sun gear to carrier assemblyand is further operable in its low-range position to couple said ringgear to a stationary member and release said first sun gear for rotationrelative to said carrier assembly; and a shift mechanism for moving saidrange sleeve between its high-range and low-range positions.
 2. Thetransfer case of claim 1 wherein said first sun gear is fixed to acoupling ring that is splined for rotation with and axial movementrelative to said input shaft in response to movement of said rangesleeve.
 3. The transfer case of claim 2 wherein said coupling ringincludes a set of clutch teeth that are releaseably engageable withclutch teeth formed on a carrier ring of said carrier assembly when saidrange sleeve is in its high-range position.
 4. The transfer case ofclaim 3 wherein said range clutch includes a clutch hub that isrotatably supported on said coupling ring and has clutch teeth inconstant meshed engagement with said clutch teeth of said coupling ring,and further comprising a synchronizer that is operably disposed betweensaid clutch hub and said carrier ring.
 5. The transfer case of claim 1wherein said range clutch includes a clutch plate fixed to saidstationary member having clutch teeth adapted for engagement with clutchteeth on said range sleeve when said range sleeve is in its low-rangeposition.
 6. The transfer case of claim 5 further comprising asynchronizer operably disposed between said clutch plate and said rangesleeve.
 7. The transfer case of claim 6 wherein said range sleeve isfurther operable in a neutral position whereat said coupling ring isuncoupled from said input shaft, said ring gear is uncoupled from saidstationary member, and said first sun gear is uncoupled from saidcarrier assembly.
 8. A transfer case comprising: an input shaft; firstand second output shafts; a first planetary gearset including a firstsun gear, a ring gear, and a first planet gear rotatably supported by acarrier and meshed with said first sun gear and said ring gear; a secondplanetary gearset including a second sun gear driving said first outputshaft, a third sun gear driving said second output shaft, a secondplanet gear rotatably supported by said carrier and meshed with saidsecond sun gear, and a third planet gear rotatably supported by saidcarrier and meshed with said second sun gear and said second planetgear; a first coupling fixed for rotation with said first sun gear; asecond coupling fixed for rotation with said ring gear, said secondcoupling moveable between a first position and a second position; athird coupling interconnecting said first coupling member for commonmovement with said second coupling member while permitting relativerotation therebetween; and a shift mechanism for moving said secondcoupling between its first and second positions, said second couplingoperable in its first position to couple said first coupling to saidinput shaft and said third coupling to said carrier, said secondcoupling is further operable in its second position to couple said firstcoupling to said input shaft and couple said ring gear to a non-rotarymember.
 9. The transfer case of claim 8 wherein said second coupling isfurther moveable to a third position whereat said first coupling isuncoupled from said input shaft and said third coupling in uncoupledfrom said carrier and said second coupling is uncoupled from saidnon-rotary member.
 10. The transfer case of claim 8 wherein said inputshaft has first and second splines and said first coupling has first andsecond splines, wherein movement of said second coupling to its firstposition causes corresponding movement of said first coupling such thatits first splines engage said second splines on said input shaft, andwherein movement of said second coupling to its second positions causessaid first coupling to move to a position whereat its second splinesengage said first splines on said input shaft.
 11. The transfer case ofclaim 8 wherein said third coupling includes clutch teeth that arereleaseably engageable with clutch teeth formed on said carrier whensaid second coupling is in its first position.
 12. The transfer case ofclaim 11 further comprising a synchronizer operably disposed betweensaid clutch teeth of said third coupling and said clutch teeth on saidcarrier.
 13. The transfer case of claim 8 wherein said second couplinghas clutch teeth operable to move into meshed engagement with clutchteeth on said non-rotary member when said second coupling is located inits second position.
 14. The transfer case of claim 13 furthercomprising a synchronizer operably disposed between said clutch teeth onsaid second coupling and said clutch teeth on said non-rotary member.15. The transfer case of claim 8 further comprising a mode clutch forcontrolling speed differentiation and torque biasing between said firstand second output shafts.
 16. The transfer case of claim 15 wherein saidmode clutch includes a friction clutch assembly operably disposedbetween said first and second output shafts, and a power-operated clutchactuator for adaptively controlling the magnitude of a clutch engagementforce exerted on said friction clutch assembly.
 17. The transfer case ofclaim 16 wherein said power-operated clutch actuator includes a cammember, an electric motor for moving said cam member, a thrust mechanismfor applying said clutch engagement force in response to movement ofsaid cam member, said electric motor receiving control signals from acontroller based on sensor signals from at least one sensor measuring aparticular operating characteristic.